DE102010060955A1 - Oscillation testing device for mechanical application of e.g. crankshaft of internal combustion engine with mechanical, dynamic test load, has control device varying phase position of unbalance force relative to another unbalance force - Google Patents

Abstract

The device (1) has an actuator (5) including unbalance exciters with respective unbalance masses rotating in opposite directions with a frequency for producing respective resultant unbalance forces. The exciters are oriented so that the unbalance forces possess a same sense of direction. A control device (2) predetermines and varies a relative phase position of one of the unbalance forces relative to the other unbalance force. The actuator is coupled with an oscillation system (9) e.g. translational oscillation system, for oscillation excitation.

Description

TECHNICAL FIELD OF THE INVENTION

The invention relates to a vibration test device by means of which any test object can be exposed to mechanical dynamic test loads. Furthermore, the invention relates to the use of such a Schwingungsprüfeinrichtung for a particular application.

STATE OF THE ART

Vibration test actuators are used for the dynamic testing of a test object. In the case of the relevant actuator, a so-called unbalance exciter is used to generate a force. In such Unwuchterregern two imbalance masses at the same speed or angular velocity, but are driven in opposite directions. The rotating imbalance masses each generate a rotating inertia or unbalance force. As a result of the opposing drive of the imbalance masses unbalance force components which have a sense of direction, which are oriented in the direction of the connecting line of the fulcrums of the imbalance masses lift on. On the other hand, the imbalance force components oriented transversely to the aforementioned direction (hereinafter referred to as "the direction of the resultant of the imbalance forces") add up to a resultant of the imbalance force. When driving the two imbalance masses of the unbalance exciter at constant speed results in a harmonious over time course of the resulting unbalance force, which can then be used as the output or test force of the actuator. When driving the imbalance masses with non-constant speed different courses of the resulting inertial force can be generated by a harmonic curve. The amplitude of the resulting inertial force of the unbalance exciter depends quadratically on the angular velocity of the imbalance masses, linearly on the mass of the imbalance masses and on the effective radius of the imbalance masses. For a given frequency to be generated of the resulting inertial force of the unbalance exciter, the amplitude of the resulting inertial force can only be altered by changing the mass of the imbalance masses, for example by supplementing or eliminating additional masses and / or changing the effective radius of the imbalance masses. Both of the aforementioned measures for changing the unbalance force of the unbalance exciter cause structural changes of the unbalance exciter or require complex automated adjustment mechanisms, for example, for the effective radius of the imbalance masses. State of the art for unbalance exciters with two oppositely rotating at the same angular velocity in synchronized alignment imbalance masses is, for example DE 23 61 183 B2 . DE 20 2007 018 292 U1 known.

It is possible that the imbalance masses are forcibly synchronized, which can be done for example in electronic or mechanical manner, for example via a gear transmission or a chain, a toothed belt or the like. This forced synchronization must be very accurate, which is why, for example DE 41 30 231 A1 and DE 42 25 564 C1 Proposals are known to allow a sensitive adjustment of the synchronous but opposite rotation of the imbalance masses.

From the field of resonance testing resonance testing machines are known in which, even with the use of an unbalance exciter, a vibration system at the or a resonance frequency thereof is excited to vibrate. An example of an application of a vibration test actuator with an unbalance exciter in a resonance testing machine is, for example DE 103 08 094 A1 known.

While usually the axes of rotation of the two imbalance masses in an imbalance exciter spaced apart from each other are oriented, disclosed EP 1 527 826 A1 a coaxial arrangement of the two imbalance masses, including an unbalanced mass as a rotating shaft with additional mass and the other unbalanced mass are arranged as a coaxial with the aforementioned shaft arranged hollow shaft with additional mass. The two additional masses mentioned are arranged offset by 180 ° to each other in the circumferential direction and dimensioned such that the same amounts of inertial forces arise for both imbalance masses despite the different distances of the additional masses of the common axis of rotation. The two imbalance masses are supported on relatively translationally guided components, so that with the guided movement, the rotation axes from the coaxial alignment can move to a parallel, spaced orientation. A synchronization of the two imbalance masses can be done according to this document via synchronized controllable motors. It is also proposed that the imbalance masses are driven by a common central drive shaft with a stationary axis via a scissor-like apart or mutually moving gear. Furthermore, it is proposed to couple the shaft and the hollow shaft in the direction of rotation, so that only one of the imbalance masses must be driven.

DE-OS 1 573 932 discloses the use of an unbalance exciter for a so-called "vibrating table". To the specimen of the vibrating table optionally oscillations with rectilinear To allow vibrational axis, circular oscillatory motion, elliptical, progressive helical or otherwise arbitrary vibration training, the document proposes to drive the imbalance masses independently via two single-acting asynchronous motors. Here, the unbalance masses should be able to be pivoted independently of each other about a common geometric axis, which is perpendicular to the axes of rotation of the imbalance masses.

On the other hand, the invention relates to vibration testing apparatuses classified in the IPC class G01M007, which relates to vibration excitation test stands in a plurality of directions for multi-dimensional vibration testing of structural or machine parts. Such Schwingungsprüfeinrichtungen serve, for example, the fatigue test, fatigue test and life test for a test specimen, which occurring at the Schwingungsprüfeinrichtung in a real operation of the specimen occurring loads are simulated or the specimen is exposed to a predetermined load program.

DE 20 2005 011 615 U1 discloses a vibration test device with which motor vehicle components, namely motors, axles, a chassis or a suspension to be tested in terms of their life or failure probability under realistic conditions as possible. For this purpose, loads acting on test drives, namely forces and movements via force and motion sensors are recorded. The detected forces and movements are then simulated or "traced" in the vibration testing device. The Schwingungsprüfeinrichtung has a kind of table against which the specimen is resiliently supported by springs with multiple degrees of freedom. The table is moved into a three-dimensional movement via several actuators, here hydraulic cylinders. A control of the hydraulic cylinder takes place via a control device according to the force and motion characteristics obtained during test drives. As a result of the elastic support of the test object relative to the table, a complex three-dimensional movement behavior of the test object results, which is detected by a movement detection device, in this case a stereo camera.

DE 691 31 185 T2 discloses a vibration test device, also with a table, in which a multiaxial vibration of the table and thus of the test object mounted on the table is to be generated. Here, a honeycomb material is used, which is to convert a one-dimensional excitation into a multiaxial excitation.

DE 35 38 229 C2 describes in the introduction to an essay "Control and exploitation of vibration as a design task" in the VDI journal, year 100 (1959), no. 25, p. 1230, Figure 20, in which a vibration tester is disclosed in which a table on two different ways can be resiliently supported, namely on the one hand as a horizontal vibration system by the table is supported by vertically aligned leaf springs, and on the other hand as a vertical vibration system by the table is supported by knee joint leaf springs. Depending on the support, the table can be excited by means of a single actuator either in horizontal oscillations or in vertical oscillations with arbitrary frequency and adjustable amplitude. The publication DE 35 38 229 C2 Against this background, the object is to modify such Schwingungsprüfeinrichtung so that the table can perform variable superimposed vertical and horizontal vibrations. For this purpose, the table is mounted both horizontally and vertically resilient and excited by an actuator in the horizontal direction and another actuator in the vertical direction with adjustable or adjustable in operation frequencies, amplitudes and phase angles to vibrations. By appropriate control of the actuators, the superimposed vertical vibrations and horizontal vibrations can be changed in any way. The actuators used are pneumatic, hydraulic or electric vibrators. It is also possible to use a crank mechanism for a Federfußpunkterregung, the crank arm of the crank mechanism should be designed to be adjustable.

DE 29 13 681 C2 also relates to a vibrating table by means of which multiaxial quasi-random oscillations in a wide frequency band are to be produced at low cost.

The non-prepublished patent application DE 10 2010 027 897.1 The applicant relates to a vibration tester for stressing a test specimen with two superimposed dynamic test loads.

OBJECT OF THE INVENTION

• – der zu erzeugenden mechanischen Beaufschlagung eines Prüflings,
• – des Energieeinsatzes für den die Belastung des Prüflings erzeugenden Aktuator und/oder
• – hinsichtlich der Regelungsmöglichkeiten für die erzeugte Belastung des Prüflings

verbessert ist.The invention has for its object to provide a Schwingungsprüfeinrichtung, which respect

• The mechanical loading of a test object to be generated,
• The use of energy for the actuator generating the load on the device under test and / or
• - Regarding the control possibilities for the generated load of the test object

is improved.

SOLUTION

The object of the invention is achieved with the features of independent claim 1. Further embodiments of this solution will become apparent according to the features of the dependent claims 2 to 19. A further solution of the problem underlying the invention is given by a use of a Schwingungsprüfeinrichtung according to claim 20th

DESCRIPTION OF THE INVENTION

According to the vibration testing device is equipped with a Resonanzprüfaktuator. This is understood to mean a test actuator in which a vibration system is coupled to an actuator in such a way that the vibration system is excited by it to oscillate. The Prüfaktuator is then coupled in any way with the DUT such that takes place according to the vibration state of the vibration system directly or indirectly, the mechanical loading of the test specimen with a periodic oscillating load. The resonance of the vibration system, for which with small excitation by the actuator a large effect, namely a large mechanical loading of the test piece can be generated, is utilized according to the invention with the Resonanzprüfaktuator used. For this purpose, the actuator of the vibration system is excited to vibrations with an excitation frequency corresponding to the resonant frequency of the vibration system. In this case, "excitation frequency corresponding excitation frequency" is understood to mean any excitation frequency which utilizes an increase in resonance of the oscillation system, in particular in the region of an amplification function of the oscillation system with an increase greater than 2, 3, 5, 10 or even more than 20, 50 or 100. Thus, the inventive equipment of Schwingungsprüfeinrichtung with a Resonanzprüfaktuator leads to a particularly energy-efficient generation of mechanical loading of the test specimen.

According to the invention but not in the Resonanzprüfaktuator any actuator such as an electromagnetic shaker, a hydraulic pulser u. Ä. use. Rather, a first unbalance exciter is used with two unbalanced masses which can be driven in opposite directions at a rotational speed and which generates a first resulting imbalance force. In addition, deviating from the prior art mentioned above, a second unbalance exciter is used which has two further imbalance masses which rotate in opposite directions at the same rotational speed as the imbalance masses of the first imbalance exciter. Here, the first unbalance exciter and the second unbalance exciter are spatially aligned in such a way that the first resulting imbalance force and the second resulting imbalance force have the same sense of direction and in particular are oriented coaxially with one another.

According to the invention, a control device is also provided, via which the relative phase position of the first resulting imbalance force relative to the second resulting imbalance force can be predetermined and changed. In the context of the present invention, both unbalance exciters may have the same or different imbalance masses and / or the same or different distances of the imbalance masses from the axes of rotation.

As explained above, when using only an unbalance exciter for a given excitation frequency an excitation force amplitude for the Schwingungsprüfsystem and the vibration system could only be changed by changing the distance of the imbalance masses from the axis of rotation and / or by changing the imbalance masses, which is a constructive intervention or a conversion of the actuator or Resonanzprüfaktuators would require. Here, the use according to the invention of two unbalance exciters provides improved control or regulation possibilities for specification of the exciter force amplitude. This will be explained on the basis of two extreme considerations: If the control device gives a phase shift of 180 ° between the two resulting imbalance forces of the two unbalance exciters, they cancel each other out (for equal distances of the imbalance masses from the axes of rotation and the same imbalance masses), so that the total Unbalance force and thus the effective excitation force amplitude despite rotating unbalance exciter 0. If, however, the phase shift of the two resulting imbalance forces 0 °, the amounts of the two resulting imbalance forces add up to a total non-zero total imbalance force whose amount corresponds to the sum of the amounts of the unbalanced forces of the two imbalance exciter and which is maximum in this case. For phase positions between the extreme phase positions of 0 ° and 180 °, the two resulting imbalance forces overlap vectorially, so that, depending on the phase position, the excitation force amplitude between zero and the maximum of the total imbalance force can be specified. (The same applies to different imbalance masses of the two imbalance exciter and / or different distances of the imbalance masses from the axes of rotation: in this case, the exciter force amplitude within a range between a minimum of the total imbalance force not equal to zero, which corresponds to the difference of the amounts of the two resulting imbalance forces, and a maximum of the total Unbalance force, which is the sum of the amounts of the two resulting unbalance forces, can be specified). Thus, the inventive design (with energy-efficient use of a Resonanzprüfaktuators) allows improved controllability for the mechanical loading of the test specimen.

It is also possible that a "run-up" of the Resonanzprüfaktuators can be improved by the inventive design. For example, the control device can specify the relative phase angle of the two unbalance exciters during run-up to the target test frequency for the rotation of the imbalance masses such that an exciting force amplitude of zero results during run-up. If the target frequency for the unbalance exciter is reached, then only the phase position of the two resulting imbalance forces must be changed so that the desired exciter force amplitude results. This is done by short-term slight change in the speeds of the drive unit for the two unbalance exciter, so that the phase position changes. For such an embodiment of the "run-up", the occurrence of transient unwanted vibrations and their unintended influence on the loading of the test object can be reduced.

It is possible that the two unbalance exciters form a common structural unit, which can thus be coupled by means of a common interface to the vibration system. Only the total imbalance force resulting from the two resulting imbalance forces is then transferred via this interface, while the resulting imbalance forces themselves form internal forces in the structural unit. It is also possible that the vibration system is integrated into the unit. In this case, the unit is coupled via an interface with the Schwingungsprüfeinrichtung or the DUT, in which case over the interface via the resonant vibrating vibration system amplified test force can be transferred.

In an alternative embodiment of the invention, the two unbalance exciters are each formed as a single units. These can then be coupled to each other via an arbitrarily designed coupling element. Such a configuration allows the use of unbalance exciters each as a single units without simultaneous use of the other unbalance exciter and also the use of two unbalance exciters together with the previously explained advantages of improved controllability of the exciter force amplitude of the two unbalance exciters.

There are many possibilities for driving the imbalance masses of the two unbalance exciters. To name just a few examples, the imbalance masses of the first unbalance exciter are drivable by a common first drive unit, while the imbalance masses of the second imbalance exciter are driven by a common second drive unit. In this case, a common control of the first and the second drive unit can then take place by a control device, which is then responsible for the specification and change in the relative phase position between the first resulting unbalance force and the second resulting unbalance force. In an alternative embodiment of the invention, each of the four imbalance masses is driven in each case by a drive unit, in which case a common control device controls the drive units of the four imbalance masses. In this case, it is also conceivable that the drive units themselves have a sub-control device, in which case one of the sub-control devices or an additional control device can be responsible for specifying and changing the relative phase position of the resulting unbalance forces. The individual mentioned control devices can also be networked with each other, in particular for the purpose of synchronization and specification of the phase position.

It is possible that the actuator formed with the imbalance exciters works in the workspace of other conventional actuators. For example, the actuator can replace a hydraulic pulser, which operates, for example, in a frequency range up to 60 Hz. However, while physically or in view of the required construction of the frequency range for excitation frequencies of other actuators is limited, the feasible excitation frequency depends on the invention formed with unbalance exciters only from the possible drive speeds for the imbalance masses, depending on the design of the drive unit used and possibly used transmissions and the positional units used for the storage of imbalance masses is almost unlimited. Accordingly, the invention proposes that in a particular embodiment of the invention, the drive speed of the imbalance masses is more than 3000 U / min, in particular more than 8000 U / min. Such high input speeds and resulting high frequencies can be ensured even with relatively small space for drive unit, actuator, Resonanzprüfaktuator and ultimately Schwingungsprüfeinrichtung.

It is quite possible that the Resonanzprüfaktuator is coupled via a suitable interface with the DUT or a support of the DUT. For a particularly compact Embodiment of the Schwingungsprüfeinrichtung the specimen itself is integrated into the vibration system, so that in particular a mass and / or rigidity of the vibration system is formed by the candidate itself.

It is entirely possible that in the Schwingungsprüfeinrichtung only one Resonanzprüfaktuator is used with a single periodic oscillating load that we generated by this Resonanzprüfaktuator. In a further development of the vibration testing device according to the invention, however, there is no one-dimensional or simple loading of the test object. Rather, according to the invention, a first periodically oscillating load and a second periodically oscillating load are used to load the test piece. The second periodically oscillating load is superimposed on the first periodically oscillating load, the two loads having different effects. For example, the loads with different effects can be forces that have different directions. It is also possible that the loads are moments with different axes of action, for example a torsional moment and / or a bending moment. It is also conceivable that one load is an arbitrary force and the other load is an arbitrary moment. Preferably, the different effect in the generation of different stress states in the specimen, for example in a shear stress on the one hand and a normal stress on the other hand, but also the superposition of similar voltages of different directions is possible.

According to the invention, the first periodically oscillating load is generated via a first resonance test actuator. This has a first vibration subsystem, which has a first resonance frequency and is excited via a first actuator, which is formed with two unbalance exciters, to oscillations at the first resonant frequency. Accordingly, a second Resonanzprüfaktuator is formed with a second vibration subsystem, which is excited via a second actuator with two unbalance exciters to resonant vibrations. The use of two Resonanzprüfaktuatoren invention has the consequence that as a result of exploiting the resonance peaks of the transfer functions of the two vibration subsystems with minimal energy input, so minimal applied by the actuators energies, large vibration amplitudes, including large loads can be generated. The utilization of the resonance thus leads to a kind of reinforcement and increase the efficiency of the Schwingungsprüfeinrichtung.

For the above-described Schwingungsprüfeinrichtungen according to the prior art vibration systems are formed with the table and springs, while the specimen is rigidly coupled to the table and is only exposed to the vibrations of the table, so that the load of the specimen is that the specimen Movement profile of the table must follow. Internal forces of the vibration system formed with the table, so for example a force in a spring supporting the table, do not act on the test specimen. On the other hand, the resonant excitation of such a vibrating table vibrator requires that the relatively large mass of the table, which in this case is part of the vibratory system, be moved by the actuator. The present invention is based on a different approach: It is entirely possible that, according to the invention, the test specimen is also supported on a kind of oscillating table or an oscillating mass. However, according to the invention, in addition to the table or the support, the first vibration subsystem and the second vibration subsystem are provided, in which the (generalized) mass or a common (generalized) mass can perform relative oscillations to the table or support. In these vibration subsystems, the dimensioning and vibration excitation can then basically be independent of the mechanical properties of the table itself and its support to the environment, whereby, of course, some interaction of the movement of the table or the support with the movement of the vibration systems may be present when the table or form the support of a vibrating chain, in which the table or the support is coupled via spring elements to the environment and in turn the specimen and the vibration systems are coupled to the table or the support.

It is entirely possible that the resonance frequencies and thus also the excitation frequencies of the actuators are different, which has the consequence that the loads generated with the Resonanzprüfaktuatoren have different frequencies. For example, due to the different excitation frequencies, maxima of the two Resonanzprüfaktuatoren occur simultaneously at a time, while they may have different time intervals from each other at different times or a maximum of Resonanzprüfaktuators can coincide with a minimum of the other Resonanzprüfaktuators.

Another embodiment of the invention uses the knowledge that it is not always advantageous if the resonance frequencies of the two vibration subsystems are different, so that for the resonance excitation of the two vibration subsystems, the actuators would have to work at different exciter frequencies. The reason for this is that for the different excitation frequencies the system responses would lead to a constantly changing superposition of the two loads, for example in the manner of a "beat", so that there are no defined test conditions. On the basis of this knowledge, the invention proposes to design and design the two vibration systems such that the resonance frequencies of the two vibration subsystems coincide (approximately). When operating both vibration subsystems with the same excitation frequency and in resonance, the loads of the specimen are very different depending on the phase position of the two oscillations of the vibration systems, which can be controlled or controlled by a control device by appropriate specification of the phase position of the resulting imbalance forces of the two imbalance exciter. If the maxima of the two vibration subsystems occur simultaneously, this leads to a higher load than if one vibration subsystem reaches the state of maximum load thereat while the other vibration subsystem has just reached a state in which the load caused by it has a zero crossing. According to the invention, this finding is ensured by the fact that a control device, which may also consist of two control units coupled to one another, actuates the first actuator and the second actuator. The control is carried out with the same exciter frequencies corresponding to the first and second resonant frequency, wherein, as already explained above, even slight deviations of the excitation frequency of the resonance frequencies are possible, provided that the resonance peak for the two vibration subsystems is utilized. Furthermore, the control is carried out by the control device such that the excitation is carried out by the two actuators with each other correlated predetermined phase position. To mention just a few examples here, the phase shift between the excitation by the two actuators can be 0 ° or 90 ° or 180 ° or any other angle between 0 ° and 180 °. It is possible that for a test, different phase relationships correlated to each other are given for a defined test period or a kind of "sweep" of the phase position is performed.

It is possible that for this embodiment, the resonance frequencies of the vibration subsystems are exactly identical. In practice, however, these will not always coincide exactly, so that claimed "approximately" equal resonance frequencies are claimed and claimed that the excitation frequency Ω "approximately" corresponds to the resonance frequencies. Hereby also deviations of the frequencies are to be detected, which still exploit an increase in resonance of the vibration subsystems. By way of example, this may be a deviation of the excitation frequency from at least one resonance frequency and / or the resonance frequencies by plus / minus 10%, in particular plus / minus 5% or plus / minus 2%. It is also possible that a frequency is used as the excitation frequency at which the transfer functions of the vibration subsystems do not fall below 80%, 90% or 95% of the maximum at the resonance frequencies.

In principle, it is conceivable that, in particular with known mechanical properties of the test object, for example the different relevant stiffnesses and the mass and mass distribution of the test object, a structural design of the vibration test device with the vibration subsystems takes place in such a way that the resonance frequencies of the vibration subsystems sufficiently match. However, if the mechanical properties of the test specimen are not known a priori or if different test specimens are to be examined with the oscillation testing device, the invention proposes that (at least) an adjusting device be provided, via which the resonant frequency of at least one oscillation subsystem can be set. It is possible that an adjustment of the resonance frequency takes place via the change in an effective in the vibration subsystem stiffness, for example, additional stiffening elements can be used in the vibration subsystem, which are thus turned on in parallel or series connection to the DUT in the power flow into the vibration subsystem. Preferably, however, an influencing of the resonant frequency of the vibration subsystem via a change in the oscillating mass. It is possible here that a mass is increased or decreased by, for example, the adjustment allows the screwing of additional masses. It is also possible that the adjustment changed the position or location of a mass, whereby a mass distribution of the vibration subsystem is changed with this brought about a change in mass or moment of inertia o. Ä. It is possible that via the adjustment a mass or rigidity in stages changeable is, as is the case for the attachment of additional mass or an additional stiffness. A particularly good adjustment is obtained when the adjustment infinitely allows the change of a location of a mass, such as the distance of an oscillating mass from a movement axis, with a sensitive adjustment via a suitable adjustment such as a spindle drive o.

It is entirely possible that the two vibration subsystems are "additionally" formed to the DUT and are coupled to the DUT, in particular via a spring element of the vibration subsystems, which then transmits the load to the DUT. For a particular embodiment according to the invention, however, the power flow in the first vibration subsystem and / or the second vibration subsystem extends over the test specimen. Thus, the specimen is part of the first and / or second vibration subsystem. As a result, oscillating internal force quantities of the vibration subsystems form the load of the test object. On the other hand, this configuration has the consequence that the rigidity of the first vibration subsystem and the second vibration subsystem is dependent on the DUT. This may be an arbitrary dependency - for example, the stiffness of the vibration subsystems of a longitudinal stiffness, transverse stiffness, bending stiffness and / or torsional stiffness of the specimen dependent.

The vibration subsystems may be of any desired design, for example as linear oscillators, flexural vibrators, torsional oscillators or rotary bending oscillators with an arbitrary number of degrees of freedom. In a particular embodiment of the invention, the first vibration subsystem is designed as a torsional vibration system in which, for example, the specimen is acted upon in the direction of its longitudinal axis as a kind of torsion spring. By contrast, for this particular embodiment, the second vibration subsystem is designed as a bending vibration system in which the specimen can be subjected to bending transverse to its longitudinal axis. For such a configuration, the torsion stress and shear stress caused by the torsion, which becomes larger radially outward from the longitudinal axis, thus comes to overlap with a normal stress, which is transverse to the longitudinal axis and transverse to the bending axis, starting from a neutral fiber of the test specimen increases to the outside. By suitable control of the phase position of the excitation of the two vibration subsystems, the specification of the temporal superposition of said loads can be specified via the control device. Quite possible in this case is that a single component forms both the torsional oscillations about an oscillating mass of the torsional vibration system that is executing about a first axis and also forms the bending vibrations about a second axis of the bending vibration system.

In a further refinement of this solution, the torsional vibration system is formed with two masses, for example two torsion discs, which perform oscillation oscillations. In the torsional vibration system, these two masses or disks are coupled together via the torsionally stressed test specimen. For such a configuration, for example, an adjustment of the resonant frequency of the torsional vibration system can be carried out by at least one oscillating rotating mass is changed by an additional mass is coupled or the distance of a partial mass of the torsion axis is changed to the mass. To form the bending vibration system acts between the two rotational vibrations about the torsion axis exporting masses or torsion disks away from a longitudinal axis or the torsion axis of the specimen, an actuator which generates a load as a force which is oriented parallel to a longitudinal axis or torsion axis of the specimen and a bending moment on the test specimen applies. Thus, the masses can also provide the lever arm for the generation of the bending moment available, so that they are multifunctional, resulting in a compact Schwingungsprüfeinrichtung. It is understood that the masses of the torsional vibration system can deviate from the disc-shaped bodies have any geometry, for example in the form of a bar or with crossing struts. About additional masses on the struts can then be done on the one hand to influence the resonant frequency of the torsional vibration system and on the other hand, the bending-vibration system.

In a further possible embodiment of the invention, the vibration subsystems are not designed as a torsional vibration system and bending vibration system. Rather, with this embodiment, the two vibration subsystems are each formed as a vibration subsystems with translational degree of freedom with vibration axes which are offset from one another or are inclined at an angle α to each other. For this embodiment of the Schwingungsprüfeinrichtung preferably come in the sample to each other inclined or offset translational oscillating forces and / or bending moments generated thereby for superimposition, which may come very close to the actual loads during operation for some candidates.

The possible applications of such a vibration testing device can be expanded if the angle of inclination of the vibration axes can be changed, so that the vibration axes can be adjusted in accordance with the loads in real operation of the test object.

For a particular inventive solution finds a Schwingungsprüfeinrichtung the type described above use for the examination of crankshafts, crankshaft bearings or Crankshaft housings of a vehicle. Examination of such specimens is prior to the present invention primarily carried out by resonant vibration subsystems in which the specimen was subjected exclusively to a load. For example, crankshafts were tested in a torsional vibration system in which the crankshaft was interposed as a continuous torsion spring between two disc-shaped masses or beams excited relative to each other to torsional vibrations. Alternatively, an examination of crankshafts was carried out by coupling the crankshaft at its end with transverse bars oriented transversely to the longitudinal axis of the crankshaft. Between the transverse beams, according to this prior art, an electromagnetic actuator has generated an oscillating force oriented parallel to the longitudinal axis of the crankshaft, which has generated an oscillating bending moment in the crankshaft. Also used was the biaxial test by means of suitable hydraulic actuators. The invention has recognized that it is not sufficient for a realistic test of crankshafts to perform in the vibration tester "somehow" an overlay of a bending moment and a torsional moment. Rather, as a result of the meandering course of the crankshaft, it has a moment of inertia of area which is relevant for the flexural rigidity and bending load and which changes greatly depending on the angular position of the crankshaft about the torsional or rotational axis. For a realistic test of the crankshaft, it is thus desirable to introduce the maximum bending moment for an angular position of the crankshaft in this, which corresponds to the angular position for which in real operation of the crankshaft, a maximum force is exerted by the connecting rod on the crankshaft. This can be done by the inventive control of the phase position of the excitation of the vibration subsystems. The above applies not only to the testing of the crankshaft itself - the same applies to the testing of crankshaft bearings or a test of a crankshaft housing or by the crankshaft bearings supporting housing elements.

Advantageous developments of the invention will become apparent from the claims, the description and the drawings. The advantages of features and of combinations of several features mentioned in the introduction to the description are merely exemplary and can come into effect alternatively or cumulatively, without the advantages having to be achieved by embodiments according to the invention. Further features are the drawings - in particular the illustrated geometries and the relative dimensions of several components to each other and their relative arrangement and operative connection - refer. The combination of features of different embodiments of the invention or of features of different claims is also possible deviating from the chosen relationships of the claims and is hereby stimulated. This also applies to those features which are shown in separate drawings or are mentioned in their description. These features can also be combined with features of different claims. Likewise, in the claims listed features for further embodiments of the invention can be omitted.

BRIEF DESCRIPTION OF THE FIGURES

In the following the invention will be further explained and described with reference to preferred embodiments shown in the figures.

1 schematically shows an operating principle of a vibration test device according to the invention with a vibration (-teil) system and an actuator.

2 shows a front view of a structural design of a Schwingungsprüfeinrichtung invention with a vibration (partial) system and an actuator.

3 shows a schematic partial section III-III through an actuator of the Schwingungsprüfeinrichtung according to 2 ,

4 schematically shows an operating principle of a vibration test device according to the invention with two vibration subsystems and two actuators.

5 shows a front view of a structural design of a Schwingungsprüfeinrichtung invention with two vibration subsystems and two actuators.

6 shows the Schwingungsprüfeinrichtung according to 5 in a top view.

7 shows that of the actuators for the embodiment according to 5 and 6 generated excitation signals with a predetermined phase shift.

8th shows the transfer functions of the two vibration subsystems for the embodiment according to 5 and 6 with resonance peaks occurring at the same resonance frequencies.

9 shows the excitation force curve of the first vibration subsystem in response to the excitation force curve of the second vibration subsystem for the embodiment according to 5 and 6 ,

10 shows a further embodiment of a Schwingungsprüfeinrichtung invention with two vibration subsystems and two actuators.

11 shows the transfer functions of the two vibration subsystems according to

10 with resonance peaks that occur at different resonant frequencies.

DESCRIPTION OF THE FIGURES

1 schematically shows a vibration tester 1 , This has a control device 2 , which via a control signal 3 an actuator 5 controls. The actuator 5 is designed with two unbalance exciters in which two imbalance masses rotate in opposite directions with the same frequency. The actuator 5 creates a total imbalance 7 , With the total imbalance 7 becomes a vibration (partial) system 9 excited to foreign excited vibrations, which, in particular after the decay of transient oscillations in the steady state, to a load 11 leads. With the burden 11 becomes a candidate 13 applied. The examinee 13 is opposite the environment 58 supported, for example, firmly clamped to a foundation or via a spring element 57 supported on a foundation. It should be noted that the schematic schematic diagram of the operating principle of the vibrational Schwingungsprüfeinrichtung 1 according to 1 not necessarily to be understood that the burden 11 a pure output of the vibration system 9 is, then the examinee 13 is supplied. Rather, the examinee 13 also part of the vibration system 9 be by this example, a spring element of the vibration system 9 forms, which with the vibrations as part of the vibration system 9 is deflected, so that the spring force the load 11 forms. For example, for the training of the vibration system 9 as a translational vibration system, a longitudinal or transverse rigidity of the specimen 13 be used as a spring element. It is also possible that a bending stiffness of the specimen 13 in a vibration system designed as a bending vibration system 9 can be used. With the actuator 5 and the associated vibration system 9 a Resonanzprüfaktuator is formed, which is the burden 11 for the examinee 13 generated.

For the embodiment according to 2 and 3 are, as far as possible, the reference numerals according to 1 used.

According to 2 and 3 is a vibration system 9 with a spring 14 formed, of which a spring base to a structural unit 15 is articulated. The construction unit 15 has a housing 16 in which there are two unbalance exciters 17 . 18 are located. Each of the unbalance exciters 17 . 18 owns a pair of imbalance masses 19 . 20 respectively. 21 . 22 , The unbalance exciter 17 generated by means of the two imbalance masses 19 . 20 a resulting imbalance 23 which for the in 3 shown position of the imbalance masses 19 . 20 oriented upwards. Horizontal inertia forces of imbalance masses 19 . 20 save each other up. On the other hand, the imbalance masses produce 21 . 22 a resulting imbalance 24 , what a 3 oriented downwards. At equal masses and distances from the axes of rotation produce the imbalance masses 19 - 22 a total imbalance 25 which results from the superimposition of the resulting imbalance forces 23 . 24 that results for the in 3 sketched phase position is zero. This is because, according to 3 the imbalance masses 19 . 20 as well as the imbalance masses 21 . 22 have a phase position, so that they rotate in opposite directions. As a result, the resulting imbalance forces 23 . 24 reach their maximum at the same time, cf. the outlined position in 3 , For the choice of a different phase relationship for the rotation of the imbalance masses 19 . 20 relative to the rotation of the imbalance masses 21 . 22 this results in a total unbalance force 25 , which is different from zero. Rotate the imbalance masses 19 . 20 such as 21 . 22 with a phase shift of zero, but in pairs in opposite directions, the result is an amount of the total unbalance force 25 , which is the sum of the amounts of the resulting unbalance forces 23 . 24 equivalent. For the in 2 illustrated embodiment carries the housing 16 also drive units for driving the imbalance masses 19 - 22 , For the illustrated embodiment, the separate drive of the imbalance masses 19 - 22 each assigned via an imbalance mass drive units 26 - 29 , The drive units 26 - 29 Become an in 2 not shown control device 2 with control signals 3 controlled, so that a defined phase position for the resulting imbalance forces 23 . 24 is given, which via a suitable control logic of the control device 2 is changeable. In the case 16 is with the imbalance masses 19 - 22 the actuator 5 educated. In the vibration system 9 is at the spring 14 a mass resiliently supported, which with the mass of the housing, the corresponding storage units, the imbalance masses 19 - 22 as well as the mass of the drive units 26 - 29 is formed. The thus formed spring-mass system, which the vibration system 9 is formed by the total imbalance 25 excited to vibrate. To the pen 14 is with the interposition of a coupling and guide element 30 an end portion of the specimen 13 coupled, the other end of the spring element 57 is supported on the environment. The examinee 13 This forms in the power flow between the spring element 57 and coupling and guide element 30 a spring element, for example, with its longitudinal or transverse stiffness depending on the mounting position of the specimen 13 and its coupling to the adjacent components. The coupling and guide element 30 is guided as shown in the vertical direction, ie parallel to the sense of direction of the unbalance forces 23 - 25 , For the illustrated embodiment, in addition to the support via the spring element 57 at the surroundings 58 , a spring-mass system formed with two vertical degrees of freedom, the rigidity of the specimen 13 and the mass of the coupling and guiding element 30 form the first spring-mass system to which the spring 14 with the mass with the housing 16 and the drive units 26 - 29 is coupled as a second spring-mass system.

4 schematically shows a vibration tester 101 , This has a control device 102 which via control signals 103 . 104 actuators 105 . 106 controls. The actuators 105 . 106 In this case, in each case, in principle, they are designed like the previously described actuator. For example, the control signals 103 . 104 to electrical signals that drive the imbalance masses in the actuators 105 . 106 correspond with the same angular velocities and given phase position.

The actuators 105 . 106 generate from the control signals 103 . 104 each total unbalance forces 107 . 108 , where it can be generalized forces, ie forces, bending moments or torsional moments in any direction, act. For the total imbalance forces 107 . 108 surrendered: F ₁ = F ₁₀ × sin (Ω ₁ t) and F ₂ = F ₂₀ × sin (Ω ₂ t + φ) with F ₁ (t) as a time-dependent total unbalance force 107 , F ₂ (t) as a time-dependent total unbalance force 108 , F ₁₀ , F ₂₀ as amplitudes of the total unbalance forces 107 . 108 , Ω ₁ = Ω ₂ = Ω as the excitation frequency given by the angular velocity of the imbalance masses and φ as the phase shift or position between the total imbalance forces 107 . 108 ,

About the controller 102 become the actuators 105 . 106 so controlled that the total unbalance forces 107 . 108 have the same excitation frequencies Ω and a phase shift φ between the total imbalance forces 107 . 108 can be specified.

With the total imbalance forces 107 . 108 each becomes a vibration subsystem 109 . 110 excited to foreign excited vibrations, which, especially after the decay of transient oscillations in the steady state, to a first load 111 through the vibration subsystem 109 caused, as well as to a second load 112 through the vibration subsystem 110 is generated leads. With these loads 111 . 112 these are internal, oscillating forces or moments of the vibration subsystems 109 . 110 or forces, which via the coupling of the vibration subsystems 109 . 110 to the examinee 113 be handed over to this. With the burdens 111 . 112 becomes the examinee 113 so that in the area of the test specimen 113 a superposition of the two loads 111 . 112 he follows. The examinee 113 is opposite the environment 158 supported, for example, firmly clamped to a foundation or via a spring element 157 supported on a foundation.

It is possible that the spring element 157 Part of a vibrating chain is, in which also the vibration subsystem 109 and / or the vibration subsystem 110 are integrated / is. It should be noted that the schematic schematic diagram of the operating principle of the Schwingungsprüfeinrichtung invention 101 according to 4 not necessarily to be understood that the burdens 111 . 112 pure output variables of the vibration subsystems 109 . 110 who are then the examinee 113 be supplied (see also the embodiment 10 ). Rather, the examinee 113 also part of at least one vibration subsystem 109 . 110 be by this example, a spring element of the vibration subsystems 109 . 110 forms, which with the vibrations as part of the vibration subsystems 109 . 110 is deflected, so that the spring force the loads 111 . 112 forms (sa the dashed and dot-dashed border in 4 as well as the embodiment 5 and 6 ). Here, different stiffnesses of the test piece 113 as a spring element of the vibration subsystems 109 . 110 be used. For example, for the formation of the vibration subsystem 109 as a translational vibration subsystem a longitudinal stiffness of the specimen 113 be used as a spring element, while a flexural rigidity of the specimen 113 in a vibration subsystem designed as a bending vibration system 110 can be used. With the actuator 105 (respectively. 106 ) and the associated vibration subsystem 109 (respectively. 110 ) is formed in each case a Resonanzprüfaktuator which the loads 111 . 112 for the examinee 113 generated.

• – über eine erste Masse 115, hier eine Platte, welche über elastische Füße oder Federelemente 116 gegenüber der Umgebung, hier einem Fundament oder Boden, abgestützt ist, und
• – eine zweite Masse 117, hier eine Drehscheibe 118, die unter Zwischenschaltung eines Torsion-Federelements 119 gegenüber der Masse 115 abgestützt ist,

mit der Umgebung gekoppelt.In the embodiments according to 5 and 6 such as 10 are the reference numerals according to 4 used for different embodiments of the invention:
For the embodiment according to 5 and 6 is the vibration tester 101 with a first vibration subsystem 109 formed, which as a torsional vibration system 114 is trained. This first vibration subsystem 109 is

• - over a first mass 115 , here a plate, which has elastic feet or spring elements 116 is supported against the environment, here a foundation or ground, and
• - a second mass 117 , here a turntable 118 , with the interposition of a torsion spring element 119 opposite to the mass 115 is supported,

coupled with the environment.

On the spring element 119 opposite side of the mass 117 relies on this a holding device 120 for the examinee 113 , here a crankshaft 121 or an axial portion of the same, from. In the holding device 120 For example, a sensor, such as a force transducer or accelerometer, or a combined torsional flexure sensor can be integrated. At the holding device 120 opposite end portion is the crankshaft 121 at another holding device 122 attached, in turn, to a mass 123 or turntable 124 is attached. The crowds 115 . 117 and 123 , the spring element 119 , the holding devices 120 . 122 as well as the examinee 113 , here the crankshaft 121 , are substantially coaxial with a longitudinal and torsion axis 125 oriented and form a torsional swinging chain. With the masses 115 . 117 and 123 a three-mass torsional vibration is formed, in which the spring elements by the spring element 119 as well as the examinee 113 , optionally with additional elasticity of the holding devices 120 . 122 , are formed. The crowds 115 . 117 and 123 can under the action of the aforementioned spring elements relative oscillating rotational movements about the longitudinal and torsional axis 125 To run. To excite such torsional vibrations acts, for example, on the mass 117 the actuator 105 , which for the illustrated embodiment for reasons of symmetry with two partial actuators 105a . 105b , which in turn each have two unbalance exciters is formed. Without the part actuators 105a . 105b basically the relative movement of the masses 115 . 117 obstruct or influence, can with a suitable control signal 103 a control device 102 via the partial actuators 105a . 105b an excitation torsional moment M _T = M _T0 sin (Ωt) at the ground 117 be generated. After transient processes have subsided, a stable periodic oscillation state results with equal or opposite rotation of the masses 115 . 117 . 123 around the longitudinal and torsion axis 125 , The part actuators 105a . 105b generate forces, which in the circumferential direction about the longitudinal and torsional axis 125 are oriented, ie vertically to the plane according to 6 , The vibration subsystem 109 is with the spring stiffness providing specimen 113 as well as the mass 123 educated. In the present case, the excitation of this vibration subsystem takes place 109 over through the actuator 105a . 105b generated torsional vibration of the mass 117 , This is the vibration subsystem 109 Component of the torsional vibration chain with the excitation by the actuator 105a . 105b in one outside the vibration subsystem 109 lying part of the Torisons vibrating chain.

The second vibration subsystem 110 is for the embodiment according to 5 and 6 formed with the roughly double-T-shaped bending torsional vibration system 126 in which the upper transverse leg of the double T is formed with the mass 123 or turntable 124 while the vertical leg of the double-T is formed with the holding device 120 , the examinee 113 and the holding device 122 and the lower transverse leg of the double-T is formed with the mass 117 or turntable 118 , When equipping the vertical leg of the double-T with a bending elasticity about a bending axis, which is vertical to the plane according to 6 This bending vibration system can be oriented 126 Perform bending oscillations, in which the transverse legs of the double-T their relative inclination to each other and the inclination relative to the horizontal in accordance with 6 change so that they swing around the previously described bending axis. It is understood that the holding devices 120 . 122 in addition to the examinee 113 contribute to the bending stiffness of the vertical leg of the T. An excitation of the bending vibration system 126 to externally excited bending vibrations carried by the actuator 106 , here for reasons of symmetry, two partial actuators 106a . 106b , each of which identifies two unbalance exciters. The part actuators 106a . 106b are at a distance 156 from the longitudinal and torsion axis 125 on opposite sides respectively at the mass 123 or turntable 124 attached. The part actuators 106a . 106b each generate total unbalance forces which are parallel to the longitudinal and torsional axis 125 are oriented. With suitable control by a control unit 102 with a control signal 104 generate the part actuators 106a . 106b Forces F (t) = F ₀ × sin (Ωt + φ). Accordingly, the excitation by the partial actuators 106a . 106b the same frequency as the excitation by the partial actuators 105a . 105b , where the excitation by the partial actuators 105a . 105b such as 106a . 106b with a phase offset of φ. The part actuators 106a . 106b are preferably designed such that they generate the said forces, but a relative rotation of the masses 117 . 123 do not obstruct or influence in the course of the forming torsional vibration. The actuators 105a . 105b such as 106a . 106b are operated simultaneously, so that in the examinee 113 Strains come to overlap, on the one hand by the actuators 105a . 105b and the torsional vibrations caused thereby and, on the other hand, the actuators 106a . 106b and the bending vibrations caused thereby are caused.

For the training of the test object 113 as a crankshaft 121 corresponds to the axis of rotation of the crankshaft 121 in operation in the installed state in the Schwingungsprüfeinrichtung 101 the longitudinal and torsion axis 125 or the vertical leg of the double-T. A crankpin 127 the crankshaft 121 which is in operation of the crankshaft 121 the connecting rod is supported, is offset laterally to the longitudinal and torsional axis 125 arranged. The longitudinal axis of the crank pin is on the connecting line of the two part actuators 106a . 106b arranged (cf. 6 ). This arrangement results in that with the bending of the vertical leg of the T of the bending vibration system 126 also cheeks 128 . 129 the crankshaft 121 be claimed on bending.

By means of the torsional vibration system 114 there is a torsional load on the crankshaft 121 which is correlated and adjusted in terms of its amplitude to a drive torque which must be transmitted during operation of the crankshaft to the other drive train. On the other hand, with the bending vibration system 126 a bending load of the crankshaft 121 simulated, which arises when in particular in the vicinity of a dead center of the crank pin 127 the working stroke of the cylinder begins with maximum or large force exertion of the connecting rod on the crank pin 127 , Thus, the illustrated arrangement of the longitudinal axis of the crank pin 127 on the connecting line of the part actuators 106a . 106b close to reality the moment of maximum bending load of the crankshaft 121 simulate. Quite conceivable, however, for example, depending on the ignition timing of the internal combustion engine, in which the crankshaft 121 to insert that is the crankshaft 121 is more or less twisted compared to the described situation.

For the illustrated embodiment, the bending resonance frequency of the bending vibration system 126 structurally fixed and is ω ₂ . On the other hand, the torsional resonance frequency is ω _{1 of} the torsional vibration system 114 changeable, including an adjustment 130 is provided. The adjustment device 130 is here with additional masses 131 . 132 formed, whose distance 133 from the longitudinal and torsion axis 125 is changeable. The additional masses 131 . 132 are from the crowd 123 or rotary plate 124 worn, so depending on the distance 133 a different mass moment of inertia of the turntable 124 results. It is possible that a change in the distance 133 about a kind of spindle shaft 134 done with opposing threaded areas, which the additional masses 131 . 132 prevail, so that with rotation of the spindle shaft 134 the masses in the direction of the longitudinal axis 125 can be moved or away from it. Additional locking or locking elements or glands can in a desired position reached then the additional masses 131 . 132 opposite the turntable 124 fix or fasten. The additional masses 131 . 132 or the adjusting device formed with these 130 is or are preferably arranged on an axis, which through the longitudinal and torsional axis 125 runs and transversely to the connection axis of the two part actuators 106a . 106b is oriented, cf. 6 , As a result, if the distance changes 133 only a change in the resonant frequency of the torsional vibration system 114 takes place, but not a change in the resonant frequency of the bending vibration system 126 takes place because the additional masses 131 . 132 undergo no or only insignificant deflections in this arrangement during the bending oscillations. By suitable adaptation of the distance 133 the additional masses 131 . 132 becomes the resonance frequency of the torsional vibration system 114 set to correspond to the resonance frequency of the bending vibration system. This is followed by an excitation of both the partial actuators 106a . 106b as well as the part actuators 105a . 105b with an excitation frequency Ω, which is the resonant frequency of both the bending vibration system 126 as well as the torsional vibration system 114 (approximately or exactly).

For the illustrated embodiment, the torsional vibration chain has a plurality of resonance frequencies. The resonance frequency used according to the invention is the resonance frequency of the first vibration subsystem 109 for which the masses 117 . 123 be rotated in opposite directions.

The crowd 117 can be used as a kind of spring-mounted table or a support for the test object 113 be regarded how this (r) is used in principle according to the prior art. However, according to the invention, the load of the specimen takes place 113 not as a result of targeted vibration stimulation of the mass 117 with vibration of the specimen and load of the specimen by its inertial forces due to the vibration, but only or by the excitation of the vibration subsystems 109 . 110 that with the examinee 113 and the crowd 123 are formed.

7 shows exemplary control signals 103 . 104 , Total imbalance 107 . 108 or load signals 111 . 112 which have a phase shift φ.

8th shows a partial section from the transfer function 135 of the vibration subsystem 109 as well as a transfer function 136 of the vibration subsystem 110 , Selected is a section in the surrounding area of the resonant peak used according to the invention. Shown here are by means of the transfer functions 135 . 136 the system responses of the vibration subsystems 109 . 110 on the vibration excitation by the actuators 105 . 106 with the same total unbalance force as a function of the exciter frequency 137 , The resonance frequencies Ω ₁ , Ω ₂ (see reference numeral 154 . 155 ) of the vibration subsystems 109 . 110 are ideally adjusted to match.

9 shows the total imbalance 107 depending on the total unbalance force 108 (or the first load 111 through the first resonance test system from the second load 112 through the second resonance test system). In the resulting ellipse, the angle between a semiaxis and the abscissa indicates the phase shift φ, while the ratio of the length of the semiaxes indicates the ratio of the amplitudes of the total unbalance forces 107 . 108 or the burdens 111 . 112 describes. Depending on the choice of exciter force amplitudes, the different, in 9 shown dashed or solid phase curves.

10 schematically shows another embodiment of a vibration tester according to the invention. Here is a candidate 113 in the form of a crankshaft 121 in a, also schematically illustrated crankcase 128 stored, with the Schwingungsprüfeinrichtung a test, preferably the crankshaft housing 138 and the support of the crankshaft 121 via the assigned bearings in this takes place. It is also possible that a bearing bracket of the crankshaft 121 the examinee 113 forms. For the illustrated embodiment are based on the crankshaft 121 , preferably at different crank pins, which are vertical to the plane according to 10 are offset, connecting rods 139 . 140 from which are oriented at an angle to each other, which corresponds to the angle of the connecting rod in the installed state of the crankshaft in a V-engine. In the crankshaft 113 remote end areas are the connecting rods 139 . 140 via a relative to the environment supported spring element 141 . 142 biased, wherein the bias voltage via biasing means 143 . 144 is changeable, which allow the displacement of a spring base. Furthermore, with the connecting rods 139 . 140 one vibration subsystem each 109 . 110 coupled, which as translational vibration subsystems 145 . 146 are formed. The vibration subsystems 145 . 146 are each with a spring element and a via the spring element to the connecting rod 139 . 140 coupled mass formed, the mass in the direction of a vibration axis 147 . 148 is guided in a sliding manner. The angle of the oscillation axes 147 . 148 corresponds to the angle α of the connecting rod 139 . 140 , ie the angle of the connecting rod in the V-engine, which in 10 with the reference number 149 is marked. Preferably, the angle 149 variable. It is possible that a connecting rod 139 with associated vibration subsystem 146 not in the circumferential direction around the crankshaft 121 slidably disposed while the other connecting rod 140 with associated vibration subsystem 146 is designed to be movable in the circumferential direction and adjustable and lockable or fastened. It is also possible that both connecting rods 139 . 140 with associated vibration subsystems 145 . 146 movable in the circumferential direction and adjustable in the Schwingungsprüfeinrichtung 101 are integrated. The drive of the vibration subsystems 145 . 146 takes place in 10 not shown way on the vibration subsystems 145 . 146 each associated actuators 105 . 106 with control via control signals 103 . 104 which a control device 102 come from, for the amplitudes, excitation frequencies and the phase shift, the above applies. The actuators 105 . 106 are each formed with two unbalance exciters. The resonance frequencies of the vibration subsystems 145 . 146 are adjustable, which can be used to control the resonance frequencies of the vibration subsystems 145 . 146 to bring into line. By means of the vibration subsystems 145 . 146 is a periodically oscillating longitudinal force with the same frequency and a given phase shift on the connecting rod 139 . 140 applied, wherein the longitudinal force in the direction of the longitudinal axes of the connecting rod 139 . 140 is oriented and over the connecting rods 139 . 140 on the crankshaft 121 as first and second load 111 . 112 be transmitted. A dimensioning of the bias of the spring elements 141 . 142 over the pretensioners 143 . 144 and the total imbalance forces that the vibration subsystems 145 . 146 to excite vibrations, can be arbitrary, so that, for example, a vibration around an average value is such that in the connecting rod 139 always a compressive force acts, so the bias voltage is greater than the amplitude of the force generated by the vibration. It is also possible that during the oscillation, the sign of the longitudinal force in the connecting rod 139 . 140 reverses, for which the amplitude of the force generated by the vibration is greater than the bias voltage.

In an embodiment of the invention, the excitation frequencies Ω of the actuators correspond 105 . 106 the matching resonant frequencies ω of the vibration subsystems 109 . 110 , In principle, use can also be made of the advantages according to the invention if the resonance frequencies of the vibration subsystems 109 . 110 do not match exactly and / or the excitation frequency of the total imbalance forces slightly different from the resonant frequency of at least one vibration subsystem 109 . 110 , The decisive factor is that an increase in resonance of the transfer functions of the vibration subsystems 109 . 110 is exploited to a sufficient extent, with small excitation energy relatively large amplitudes of the vibrations of the vibration subsystems 109 . 110 bring about. This situation is in 11 exemplified, in which the transfer functions 150 . 151 the vibration subsystems 109 . 110 are shown, wherein the resonance frequencies 154 . 155 the transfer functions 150 . 151 are slightly offset from each other, but still have the transfer functions sufficiently overlapping resonance peaks. In the 11 Dashed line shows a Resonanzüberhöhungsniveau 152 in, in the area of or above the the vibration subsystems 109 . 110 to be operated. The intersections of the dashed resonance peak 152 with the transfer functions 150 . 151 give an excitation frequency range 153 for, when choosing an excitation frequency in this excitation frequency range 153 the resonance peak at least the resonance peak level 152 is equal to or greater than this. Preferably, the resonant frequencies of the vibration subsystems are different 109 . 110 by a maximum of 10%, for example a maximum of 5% and preferably by a maximum of 2% from each other, wherein the selected excitation frequency for the two vibration subsystems 109 . 110 likewise by the abovementioned percentage ranges at most of at least one of the resonance frequencies of the vibration subsystems 109 . 110 differs.

In a modified embodiment, the vibration subsystems 109 . 110 be operated with different excitation frequencies, but each about the resonance frequencies 154 . 155 correspond. In this case, too, the resonance frequencies 154 . 155 stronger or clearer than this 11 the case is that the resonance peaks can not have overlaps either.

According to the invention, at least one exciter frequency is preferably specified without it being absolutely necessary to regulate the amplitude of the total unbalance force. However, it is quite possible that the system response of the vibration subsystems 109 . 110 is monitored and takes place to prevent damage or excessive vibration amplitudes, a regulation of the amplitude of the total unbalance force. For this purpose, displacement sensors or force sensors in the vibration subsystems 109 . 110 be integrated. In the vibration system 101 according to 5 and 6 is preferably the mass or mass moment of inertia of the masses 115 . 117 bigger than that of the crowd 123 , The actuators 106a . 106b are driven in opposite directions with a phase shift of 180 °. On the other hand, the actuators 105a . 105b applied in the same direction.

The invention is not limited to the testing of crankshafts - rather, a test of any test specimens in the form of components or material samples by means of Schwingungsprüfeinrichtung invention 1 . 101 respectively. Also, the invention is not limited to the construction of the vibration tester 101 with the vibration subsystems 109 . 110 as a bending vibration system and torsional vibration system or with two translational vibration subsystems - rather any resonant operated vibration subsystems can be used in the invention.

For the embodiment according to 10 is for example the over the spring elements 141 . 142 generated static preload 10 kN while using the vibration subsystems 145 . 146 an oscillating load with an amplitude of 20 kN is generated. It is possible that about the vibration tester 101 according to 10 a separation plane of the housing and a vertical housing wall, on which are supported bearings of the crankshaft, the screw joints between the lower and upper housing plate, a bearing bridge or a so-called Bedplate is examined. Because the actuators 105 . 106 are each formed with the two unbalance exciters, can be a sensitive adjustment of the excitation frequency and adaptation of the same to the resonant frequencies in the vibration subsystems 109 . 110 independently, independent of which an independent control of the amplitudes of the total imbalance forces can take place. In particular with respect to hydraulically pulsating actuators, the formation of the actuators 105 . 106 each with two unbalance exciters the advantage that an upper limit for the exciter frequency, as it exists for a hydraulic excitation example, 60 Hz, even without much additional effort does not exist or can be moved to higher frequencies.

In the description is only a burden of the examinee 13 . 113 described via one or two Resonanzprüfaktuatoren. In the context of the present invention, more than two Resonanzprüfaktuatoren find use for the superposition of more than two loads.

LIST OF REFERENCE NUMBERS

1

Schwingungsprüfeinrichtung

2

control device

3

control signal

5

actuator

7

Overall imbalance force

9

Vibration (partially) system

11

burden

13

examinee

14

feather

15

unit

16

casing

17

unbalance exciter

18

unbalance exciter

19

unbalanced mass

20

unbalanced mass

21

unbalanced mass

22

unbalanced mass

23

resulting imbalance

24

resulting imbalance

25

Overall imbalance force

26

power unit

27

power unit

28

power unit

29

power unit

30

Coupling and guide element

57

spring element

58

Surroundings

101

Schwingungsprüfeinrichtung

102

control device

103

control signal

104

control signal

105

actuator

106

actuator

107

Overall imbalance force

108

Overall imbalance force

109

Vibration subsystem

110

Vibration subsystem

111

first load

112

second load

113

examinee

114

Torsional vibration system

115

Dimensions

116

spring element

117

Dimensions

118

turntable

119

spring element

120

holder

121

crankshaft

122

holder

123

Dimensions

124

turntable

125

Longitudinal and torsion axis

126

Bending vibration system

127

crank pins

128

cheek

129

cheek

130

adjustment

131

additional mass

132

additional mass

133

distance

134

spindle shaft

135

transfer function

136

transfer function

137

excitation frequency

138

crankcase

139

pleuel

140

pleuel

141

spring element

142

spring element

143

biasing means

144

biasing means

145

translatory vibration subsystem

146

translatory vibration subsystem

147

axis of oscillation

148

axis of oscillation

149

angle

150

transfer function

151

transfer function

152

Resonant peak level

153

Excitation frequency range

154

resonant frequency

155

resonant frequency

156

distance

157

spring element

158

Surroundings

QUOTES INCLUDE IN THE DESCRIPTION

This list of the documents listed by the applicant has been generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• DE 2361183 B2 [0002]
• DE 202007018292 U1 [0002]
• DE 4130231 A1 [0003]
• DE 4225564 C1 [0003]
• DE 10308094 A1 [0004]
• EP 1527826 A1 [0005]
• DE 1573932A [0006]
• DE 202005011615 U1 [0008]
• DE 69131185 T2 [0009]
• DE 3538229 C2 [0010, 0010]
• DE 2913681 C2 [0011]
• DE 102010027897 [0012]

Claims (20)

Vibration tester ( 1 ; 101 ) for the mechanical loading of a test object ( 13 ; 113 ) with a periodically oscillating load ( 11 ; 111 . 112 ) with a) a Resonanzprüfaktuator with an oscillatory system ( 9 : 109 . 110 ) and one for vibrational excitation with the vibration system ( 9 ; 109 . 110 ) coupled actuator ( 5 ; 105 . 106 ), b) wherein the actuator ( 5 ; 105 . 106 ) the vibration system ( 9 ; 109 . 110 ) is excited to oscillations with an excitation frequency (Ω ₁ ), the resonance frequency (ω ₁ ) of the vibration system ( 9 ; 109 . 110 ), whereby the periodically oscillating load ( 11 ; 111 . 112 ) of the test piece ( 13 ; 113 ), characterized in that the actuator ( 5 ; 105 . 106 ) with c) a first unbalance exciter ( 17 ) with two unbalanced masses rotating in opposite directions with a frequency Ω ₁ ( 19 . 20 ), which has a first resulting unbalance force ( 23 ), and d) a second unbalance exciter ( 18 ) with two unbalanced masses rotating in opposite directions with the frequency Ω ₁ ( 21 . 22 ), which has a second resulting unbalance force ( 24 ), e) the first unbalance exciter ( 17 ) and the second unbalance exciter ( 18 ) are oriented such that the first resulting unbalance force ( 23 ) and the second resulting unbalance force ( 24 ) have the same sense of direction and f) a control device ( 2 ; 102 ) is present, via which the relative phase position of the first resulting unbalance force ( 23 ) against the second resulting unbalance force ( 24 ) can be specified and changed. Vibration tester ( 1 ; 101 ) according to claim 1, characterized in that the two unbalance exciters ( 17 . 18 ) an assembly ( 15 ) form. Vibration tester ( 1 ; 101 ) according to claim 1, characterized in that the two unbalance exciters ( 17 . 18 ) are each formed as a single units, which can be coupled to one another via a coupling element. Vibration tester ( 1 ; 101 ) according to one of the preceding claims, characterized in that a) the imbalance masses ( 19 . 20 ) of the first unbalance exciter ( 17 ) are drivable by a common first drive unit, b) the unbalanced masses ( 21 . 22 ) of the second unbalance exciter ( 18 ) are drivable by a common second drive unit and c) the control device ( 2 ; 102 ) drives both the first and the second drive unit. Vibration tester ( 1 ; 101 ) according to one of claims 1 to 3, characterized in that each of the four imbalance masses ( 19 . 20 . 21 . 22 ) each of a drive unit ( 26 . 27 . 28 . 29 ) and the control device ( 2 ; 102 ) the drive units ( 26 . 27 . 28 . 29 ) of the four imbalance masses ( 19 . 20 . 21 . 22 ). Vibration tester ( 1 ; 101 ) according to claim 5, characterized in that by means of the drive units ( 26 . 27 . 28 . 29 ) the four imbalance masses ( 19 . 20 . 21 . 22 ) with a drive speed of more than 3000 U / min, in particular of more than 8000 U / min, are drivable. Vibration tester ( 1 ; 101 ) according to one of the preceding claims, characterized in that the test object ( 13 ; 113 ) in the vibration system ( 9 ; 109 . 110 ) or with the vibration system ( 9 ; 109 . 110 ) is coupled. Vibration tester ( 101 ) according to one of the preceding claims, characterized in that the test object ( 113 ) with a first periodically oscillating load ( 111 ) and a second periodically oscillating load ( 112 ), which of the first load ( 111 ) is superimposed and has a different effect than the first load ( 112 ), with a) a first Resonanzprüfaktuator having a first vibration subsystem ( 109 ) and one for vibrational excitation with the first vibration subsystem ( 109 ) coupled first actuator ( 105 ), which is equipped with two unbalance exciters ( 117 . 118 ), wherein the first vibration subsystem ( 109 ) via the first actuator ( 105 ) is excited to oscillations with an excitation frequency (Ω ₁ ), whereby the first periodically oscillating load ( 111 ), b) a second Resonanzprüfaktuator, the second vibration subsystem ( 110 ) and one for vibrational excitation with the second vibration subsystem ( 110 ) coupled second actuator ( 106 ), which is equipped with two unbalance exciters ( 117 . 118 ), wherein the second vibration subsystem ( 110 ) via the second actuator ( 106 ) is excited to oscillations with an excitation frequency (Ω ₂ ), whereby the second periodically oscillating load ( 112 ), where c) the first vibration subsystem ( 109 ) and the second vibration subsystem ( 110 ) against the support of the test piece ( 113 ) to the environment ( 158 swing and d) a control device ( 102 ) is provided, via which the first actuator ( 105 ) and the second actuator ( 106 ). Vibration tester ( 101 ) according to claim 8, characterized in that the first actuator ( 105 ) and the second actuator ( 106 ) from a common control device ( 102 ). Vibration tester ( 101 ) according to claim 8 or 9, characterized in that a) the first vibration subsystem ( 109 ) has a first resonance frequency (ω ₁ ), b) the first vibration subsystem ( 109 ) via the first actuator ( 105 ) is excited into oscillations with an exciter frequency (Ω ₁ ) approximately at the first resonant frequency (ω _i ), c) the second oscillatory subsystem ( 110 ) has a second resonance frequency (ω ₂ ), and d) the second vibration subsystem ( 110 ) via the second actuator ( 106 ) is excited to oscillate at an excitation frequency (Ω ₂ ) approximately at the second resonant frequency (ω ₂ ). Vibration tester ( 1 ) according to claim 10, characterized in that the resonance frequency (ω ₁ ) of the first vibration subsystem ( 109 ) and the resonance frequency (ω ₂ ) of the second vibration subsystem ( 110 ) and the control device ( 102 ) is designed so that the first actuator ( 105 ) and the second actuator ( 106 ) a) with the same excitation frequencies (Ω = Ω ₁ = Ω ₂ ), which correspond to the first and second resonance frequency (ω ₁ , ω ₂ ), and b) with mutually correlated phase position (φ) are controllable. Vibration tester ( 101 ) according to one of claims 8 to 11, characterized in that the flow of force in the first vibration subsystem ( 109 ; 114 ) and / or the second vibration subsystem ( 110 ; 126 ) over the test object ( 113 ), so that the rigidity of the first vibration subsystem ( 109 ; 114 ) and / or the second vibration subsystem ( 110 ; 126 ) of the examinee ( 113 ) is dependent. Vibration tester ( 101 ) according to one of claims 10 to 12, characterized in that at least one adjusting device ( 130 ) for the first resonance frequency (ω ₁ ) of the first vibration subsystem ( 109 ; 114 ) and / or for the second resonant frequency (ω ₂ ) of the second vibration subsystem ( 110 ; 126 ) is provided. Vibration tester ( 101 ) according to claim 13, characterized in that by means of the adjusting device ( 130 ) an oscillating mass or moment of inertia of a vibration subsystem ( 109 ; 110 ) is changeable. Vibration tester ( 101 ) according to claim 13 or 14, characterized in that via the adjusting device ( 130 ) a stiffness of a vibration subsystem ( 109 ; 110 ) is changeable. Vibration tester ( 101 ) according to one of claims 8 to 15, characterized in that a) the first vibration subsystem ( 109 ) a torsional vibration subsystem ( 114 ) and b) the second vibration subsystem ( 110 ) a bending vibration subsystem ( 126 ). Vibration tester ( 101 ) according to claim 16, characterized in that a) the torsional vibration subsystem ( 114 ) with two rotational oscillations carrying masses ( 117 ; 123 ) formed over the torsion-stressed specimen ( 113 ) and b) for forming the bending vibration subsystem ( 128 ) between the two rotational oscillations exporting masses ( 117 ; 123 ) away from the longitudinal axis ( 125 ) of the test piece ( 113 ) an actuator ( 106 ), which generates a force parallel to the longitudinal axis ( 125 ) of the test piece ( 113 ) and a bending moment on the test specimen ( 113 ). Vibration tester ( 101 ) according to one of claims 8 to 15, characterized in that the first vibration subsystem ( 109 ) as well as the second vibration subsystem ( 110 ) each as a vibration subsystem ( 145 . 146 ) are formed with translational degree of freedom with vibration axes ( 147 . 148 ), which are offset from each other or inclined at an angle α to each other. Vibration tester ( 101 ) according to claim 18, characterized in that the inclination angle α of the oscillation axes ( 147 . 148 ) is changeable. Use of a vibration test device ( 101 ) according to claim 18 or 19 for the testing of crankshafts ( 121 ), Crankshaft bearings or crankshaft housings ( 138 ).

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